Heat Exchanger Comprising Wave-shaped Fins

ABSTRACT

A heat exchanger for exchanging heat between a first fluid and a second fluid is disclosed; wherein the heat exchanger has improved thermal efficiency and low fluid back pressure. The heat exchanger comprises conduits for conveying the first fluid, wherein the conduits include a plurality of flow passages. The flow passages are defined by a plurality of fins that are continuous along the direction of flow of the first fluid. Each fin includes a wave-shaped region, and adjacent fins sub-divide each flow passage into first and second sections that are interposed by a third section. The wave-shape of the fins creates a continuously varying cross-sectional area for each third section. The variation of the cross-sectional area of the third section, coupled with the wave-shape of the fins, induces a swirl flow between the third section and each of the first and second sections. This swirl flow improves the efficiency of the overall convection heat transfer in each conduit. Further, the overall cross-sectional area of each conduit remains constant even as the cross-sectional areas of individual flow passages changes, which mitigates the development of back pressure in the flow of the first fluid.

FIELD OF THE INVENTION

The present invention relates to energy conversion in general, and, more particularly, to heat exchangers.

BACKGROUND OF THE INVENTION

The Earth's oceans are continually heated by the sun and cover nearly 70% of the Earth's surface. The temperature difference between deep and shallow waters contains a vast amount of solar energy that can potentially be harnessed for human use. In fact, it is estimated that the thermal energy contained in the temperature difference between the warm ocean surface waters and deep cold waters within ±10° of the Equator represents a 3 Tera-watt (3×10¹² W) resource.

The total energy available is one or two orders of magnitude higher than other ocean-energy options such as wave power, but the small magnitude of the temperature difference makes energy extraction comparatively difficult and expensive, due to low thermal efficiency.

Ocean thermal energy conversion (“OTEC”) is a method for generating electricity which uses the temperature difference that exists between deep and shallow waters to run a heat engine. A heat engine is a thermodynamic device placed between a high temperature reservoir and a low temperature reservoir. As heat flows from one reservoir to the other, the engine converts some of the heat to work. This principle is used in steam turbines and internal combustion engines. Rather than using heat energy from the burning of fuel, OTEC power draws on temperature differences caused by the sun's warming of the ocean surface.

One heat cycle suitable for OTEC is the Rankine cycle, which uses a low-pressure turbine. Systems may be either closed-cycle or open-cycle. Closed-cycle systems use a fluid with a low boiling point, such as ammonia, to rotate the turbine to generate electricity. Warm surface seawater is pumped through a heat exchanger where the low-boiling-point fluid is vaporized. The expanding vapor turns the turbo-generator. Then, cold, deep seawater—pumped through a second heat exchanger—condenses the vapor back into a liquid, which is then recycled through the system. Open-cycle engines use the water heat source as the working fluid.

As with any heat engine, the greatest efficiency and power is produced with the largest temperature difference. This temperature difference generally increases with decreasing latitude (i.e., near the equator, in the tropics), but evaporation prevents the surface temperature from exceeding 27° C. Also, the subsurface water rarely falls below 5° C. Historically, the main technical challenge of OTEC was to generate significant amounts of power, efficiently, from this very small temperature ratio. But changes in the efficiency of modern heat exchanger designs enables performance approaching the theoretical maximum efficiency.

OTEC systems have been shown to be technically viable, but the high capital cost of these systems has thwarted commercialization. Heat exchangers are the second largest contributor to OTEC plant capital cost (the largest is the cost of the offshore moored vessel or platform). The optimization of the enormous heat exchangers that are required for an OTEC plant is therefore of great importance and can have a major impact on the economic viability of OTEC technology.

The primary function of a heat exchanger is to efficiently transfer thermal energy from one media to another. Heat exchangers of various types are widely used across many applications such as in air conditioners, industrial chemical plants and power generation. For OTEC systems, while there are many conventional heat-exchanger designs that can be considered, there are, as a practical matter, no good choices.

Almost all heat exchangers can be classified by two types of geometries: Shell and Tube Heat Exchangers or Compact and Extended Surface Heat Exchangers. Of the Compact and Extended Surface Heat Exchanger geometry, two common types are Plate-Fin Heat Exchangers and Extruded-Fin Heat Exchangers. Plate-fin heat exchangers are typically used in the Cryogenics Industry and extruded-fin heat exchangers are most commonly used in the Automotive Industry.

Conventional shell and tube heat exchangers are widely available for marine use. But the overall heat transfer coefficient, U, that is associated with reasonable pressure drops for OTEC is typically below 2000 W/m²K. This drives the size and cost for this type of heat exchanger and reduces its economic viability.

Plate-fin heat exchangers are assembled as an array of stacked aluminum plates with thin corrugated sheets of fins between the plates. The entire array is then brazed together to form a heat exchanger core. Unfortunately, plate-fin heat exchangers are undesirable for many applications because of high material and fabrication costs. Second, brazed joints are poorly suited to applications in which corrosive media are used. In OTEC applications, brazed joints are particularly susceptible to galvanic corrosion when exposed to seawater. Third, plate-fin heat exchangers are typically characterized by low thermal and/or flow efficiency. Such designs suffer from varying amounts of fluid flow resistance, and still do not eliminate manufacturing cost and corrosion issues for very large scale assemblies, however. Fourth, the small passages found in typical plate-fin heat exchangers are prone to biofouling. Fifth, maintenance, such as refitting, repair, and refurbishment, on plate-fin heat exchangers is challenging due to the difficulty of accessing their internal regions.

Conventional extruded-fin heat exchangers are made from standard aluminum extrusions with thin aluminum sections to increase thermal efficiency. Extruded-fin heat exchangers are typically much less expensive to produce than plate-fin heat exchanger cores because of the elimination of the additional fin component and brazing operations. The trade-off with this design is that they typically have lower thermal performance due to a boundary layer created in the fluid(s) being used. This boundary layer is a physical phenomenon that can only be eliminated by disrupting the fluid flow within the heat exchanger core to create stirring turbulences. This issue has been studied for decades and has been solved for various applications in different ways with limited success.

Typically, plain fins are mounted on the top and bottom of the extruded channels. These fins are much like the rectangular fins in the plate-fin heat exchanger because they are straight and uninterrupted along the full length of flow. Fins that are straight along the flow length tend to develop fluid boundary layers that are quite thick, which results in lower values of the heat transfer coefficient. A plain fin arrangement will exhibit relatively low pressure drop but have relatively low heat transfer. More complex fin designs that provide disruptions to fluid flow can improve heat transfer; however, complex fin designs suffer from a higher pressure drop through the heat exchanger.

With today's growing need for energy, using a renewable constant source is a desirable solution. As a consequence, there is a renewed interest in OTEC power plants. But development of an OTEC heat exchanger that accommodates high flow rates while minimizing pumping parasitic losses and offering long life in the ocean environment remains elusive.

SUMMARY OF THE INVENTION

The present invention provides a heat exchanger having higher heat transfer efficiency than heat exchangers known in the prior art. Heat exchangers in accordance with the present invention are particularly well-suited for use in OTEC systems.

An embodiment of the present invention comprises first conduits for conveying a first fluid through the heat exchanger and second conduits for conveying a second fluid through the heat exchanger, wherein the first fluid and second fluid are thermally coupled by the heat exchanger. The first conduits include flow passages that induce turbulence in the flowing first fluid without significantly increasing fluid back pressure. The turbulence is induced by wave-shaped fins that project into each first conduit to form a plurality of flow passages. The wave-shaped fins are continuous along the direction of fluid flow. Adjacent pairs of wave-shaped fins define three sections in each flow passage: first and second sections that are interposed by a third section. The wave-shape of the fins results in a continuous variation of the cross-sectional area of the third section along the direction of fluid flow. As this cross-sectional area changes, the first fluid flowing through the flow passage is forced to exchange between the third section and each of the two remaining sections of the flow passage. This exchange of fluid between the three sections induces a swirl, or vortex, flow in the first and second sections, which increases the overall convection heat transfer in the flow passage.

The fins are arranged within each first conduit such alternating fins project into the first conduit from opposite surfaces and so that the wave shapes of adjacent fin pairs are offset by a phase difference. This phase difference leads to a continuously periodic change on the cross-sectional area of the third section along the length of the flow passage. As the cross-sectional area shrinks, first fluid is “squeezed” from each third section into the first and second sections of each flow passage. As the cross-sectional area of the third section increases, first fluid is drawn back into the third section from the other two sections. Further, the wave-shape of the fins defines a shape of the third section that induces the first fluid to swirl as it enters and exits the first and second sections. This swirl flow creates turbulence that enhances heat transfer between the first fluid and walls of the flow passages.

It is a further aspect of the invention that the first conduits avoid inducing a significant fluidic back pressure while conveying the first fluid through the heat exchanger. Increased back pressure of the fluid is mitigated by the fact that the overall cross-sectional area of the first conduits remains the same even while the cross-sectional areas of individual flow passages within it change. The consistency of overall cross-sectional area of the conduits results from the complimentary nature of adjacent flow passages within them. Specifically, as the cross-sectional area of a first flow passage is shrinking, the cross-sectional area of its adjacent flow passages is increasing by a commensurate amount. As a result, the sum of the cross-sectional areas of all flow passages in a given first conduit remains constant.

An embodiment of the present invention comprises a heat exchanger that thermally couples a first fluid and a second fluid, wherein the heat exchanger comprises: (1) a first plate comprising a first material that is thermally conductive, wherein the first plate comprises; (i) a first conduit for conveying the first fluid along a first direction, wherein the first conduit comprises at least one first channel; and (ii) a first plurality of first fins, wherein each first fin is continuous along the first direction, and wherein each first fin comprises a first fin portion that has a first periodic shape having a first periodicity along the first direction; and (2) a second plate comprising the first material, wherein the second plate comprises; (i) a second conduit for conveying the first fluid along the first direction, wherein the second conduit comprises at least one second channel; a (ii) second plurality of second fins, wherein each second fin is continuous along the first direction, and wherein each second fin comprises a second fin portion that has the first periodic shape having the first periodicity along the first direction; wherein the first plurality of first fins and the second plurality of second fins collectively define a plurality of passages for conveying the second fluid along the first direction, and wherein each of the plurality of passages comprises a first section, second section, and third section; and wherein the cross-sectional area of the first section varies along the first direction based on the first periodicity, and wherein the variation of the cross-sectional area of the first section induces flow of the second fluid between the first section and each of the second section and third section.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 depicts a schematic diagram of an OTEC power generation system in accordance with an illustrative embodiment of the present invention.

FIG. 2A depicts a schematic drawing of a first heat exchanger in accordance with the prior art.

FIG. 2B depicts a perspective view of a plate 202.

FIG. 3A depicts a schematic drawing of a second heat exchanger in accordance with the prior art.

FIG. 3B depicts a perspective view of a section of tube 302.

FIG. 3C depicts a schematic of a portion of heat exchanger 300.

FIG. 4 depicts a perspective view of a heat exchanger in accordance with the illustrative embodiment of the present invention.

FIG. 5A depicts a top view of a portion of plate 402 prior to the formation of a wave-shape on its fins in accordance with the illustrative embodiment of the present invention.

FIG. 5B depicts a top view of a portion of plate 402, after the formation of a wave-shape on its fins.

FIG. 5C depicts a bottom view of a portion of plate 402, after the formation of a wave-shape on its fins.

FIG. 6A depicts a representational cross-sectional view of conduit 410-2 in accordance with the illustrative embodiment of the present invention.

FIG. 6B depicts a top view, through line a-a of FIG. 6A, of conduit 410-2 in accordance with the illustrative embodiment of the present invention.

FIG. 6C depicts a detailed cross-sectional view of a portion of conduit 410-2.

FIG. 7 provides operations of a method for thermally coupling a first fluid and a second fluid in accordance with the illustrative embodiment of the present invention.

FIG. 8A depicts a flow passage 602-i within in conduit 410-2, at point A depicted in FIG. 6B.

FIG. 8B depicts a flow passage 602-i within in conduit 410-2, at point B depicted in FIG. 6B.

FIG. 8C depicts a flow passage 602-i within in conduit 410-2, at point C depicted in FIG. 6B.

FIG. 9A depicts a representational cross-sectional view through a portion of a heat exchanger at a first point along the direction of fluid flow, in accordance with an alternative embodiment of the present invention.

FIG. 9B depicts a representational cross-sectional view through a portion of a heat exchanger at a second point along the direction of fluid flow, in accordance with the alternative embodiment of the present invention.

DETAILED DESCRIPTION

FIG. 1 depicts a schematic diagram of an OTEC power generation system in accordance with an illustrative embodiment of the present invention. OTEC system 100 comprises turbogenerator 104, closed-loop conduit 106, heat exchanger 110-1, heat exchanger 110-2, pumps 114, 116, and 124, and conduits 120, 122, 128, and 130. OTEC system 100 is deployed in water body 136 wherein a suitable temperature difference exists between water near the surface and water located at a deep level of water body 136.

Turbo-generator 104 is a conventional turbine-driven generator. Turbogenerator 104 is mounted on floating platform 102, which is a conventional floating energy-plant platform. Platform 102 is anchored to the ocean floor by mooring line 132 and anchor 134, which is embedded in the ocean floor. In some instances, platform 102 is not anchored to the ocean floor but is allowed to drift. Such a system is sometimes referred to as a “grazing plant.”

In typical operation, pump 114 pumps a primary fluid (i.e., working fluid 108), in liquid form, through closed-loop conduit 106 to heat exchanger 110-1. Ammonia is often used as working fluid 108 in OTEC systems; however, it will be clear to one skilled in the art that any fluid that evaporates at the temperature of the water in surface region 118 and condenses at the temperature of the water in deep water region 126 is suitable for use as working fluid 108 (subject to material compatibility requirements).

Heat exchanger 110-1 and 110-2 are configured for operation as an evaporator and condenser, respectively. One skilled in the art will recognize that the operation of a heat exchanger as evaporator or condenser is dependent upon the manner in which it is configured within system 100. Heat exchanger 110 is described in detail below and with respect to FIGS. 4 through 7C.

In order to enable its operation as an evaporator, pump 116 draws warm secondary fluid (i.e., seawater from surface region 118) into heat exchanger 110-1 via conduit 120. At heat exchanger 110-1 heat from the warm water is absorbed by working fluid 108, which induces working fluid 108 to vaporize. After passing through heat exchanger 110-1, the warm water is ejected back into water body 136 via conduit 122. In a typical OTEC deployment, the water is surface region 118 is at a substantially constant temperature of approximately 25 degrees centigrade (subject to weather and sunlight conditions).

The expanding working fluid 108 vapor is forced through turbogenerator 104, thereby driving the turbogenerator to generate electrical energy. The generated electrical energy is provided on output cable 112. Once it has passed through turbogenerator 104, the vaporized working fluid enters heat exchanger 110-2.

At heat exchanger 110-2, pump 124 draws cold secondary fluid (i.e., seawater from deep water region 126) into heat exchanger 110-2 via conduit 128. The cold water travels through heat exchanger 110-2 where it absorbs heat from the vaporized working fluid. As a result, working fluid 108 condenses back into liquid form. After passing through heat exchanger 110-2, the cold water is ejected into water body 136 via conduit 130. Typically deep water region 126 is 1000+ meters below the surface of water body 136, at which depth water is at a substantially constant temperature of a few degrees centigrade.

Pump 114 pumps the condensed working fluid 108 back into heat exchanger 110-1 where it is again vaporized; thereby continuing the Rankine cycle that drives turbogenerator 104.

FIG. 2A depicts a schematic drawing of a first heat exchanger in accordance with the prior art. Heat exchanger 200 comprises a plurality of plates 202, and housing 208. Heat exchanger 200 is in accordance with a heat exchanger described by T. F. Brise in U.K. Patent Application GB2424265, which was published Sep. 9, 2006, and which is incorporated by reference herein.

Each plate 202 is an extruded body that comprises a plurality of channels 204, which convey a first fluid along the z-direction as shown. In some cases, each of channels 204 comprises projections directed inward from the surface of the channel. These projections increase the heat transference between the first fluid and the plate material.

Plates 202 are stacked and welded together to form heat exchanger core 210. Adjacent plates in heat exchanger core 210 collectively define conduits 212 for conveying a second fluid along the z-direction. Each plate 202 further comprises a plurality of fin segments 206, which depend from the upper and lower surfaces of the body of plate 202 and project into conduits 212.

Each fin segment 206 is a discrete element that is formed from fins that are initially continuous along the z-direction. To form fin segments 206, these initially continuous fins are “segmented and the fin segments twisted so that at least the upper parts of the fin segments are at an angle” with respect to the z-direction (i.e., the longitudinal axis of the tube). According to Brise, “Segmenting the fins and twisting the fin segments creates lateral flow paths across the surface of the tube along which coolant can flow.”

Heat exchanger core 210 is located within heat exchanger 200 via housing 208. Housing 208 locates heat exchanger core 210 by means of seats 214, which receive the uppermost and lowermost fin segments that project from the heat exchanger core. Once heat exchanger core 210 is positioned within sleeve 208, weld joints are formed at seats 214 to firmly fix the heat exchanger core in place.

FIG. 2B depicts a perspective view of a plate 202. As is readily seen from the figure, the top and bottom outer surfaces of plate 202 are covered by rows of individual fin segments 206. These fin segments are arranged along the z-direction in rows.

As the second fluid flows along the z-direction, the discontinuities between fin segments 206 create turbulence that enhances heat transfer between the second fluid and plate 202.

Unfortunately, heat exchangers such as heat exchanger 200 have several disadvantages. First, in many cases the formation of fin segments 206 shears off much of the base of each fin segment. As a result, fin segments 206 are highly susceptible to fracture during use. Fractured fins are likely to lodge in conduits 212 and restrict the flow of the second fluid through the heat exchanger.

Second, when the plates 202 are stacked together to form conduits 212, fin segments 206 are interlocked such that they nearly completely obstruct the flow path for the second fluid. As a result, fin segments 206 cause a large pressure-drop through conduits 212.

Third, it is well-known to those skilled in the art that discontinuities, such as those between fin segments 206, significantly increase back pressure.

Fourth, it is well known that discontinuities in a flow path are prone to impurity build-up, which leads to fouling during use. For example, the use of serrated fins, such as those included in heat exchanger 200, is discouraged in many heat exchanger design handbooks, such as “Heat Exchanger Design Handbook,” written by T. Kuppan. In general, it is difficult and expensive to clean fouled fluid conduits of a heat exchanger, and nearly impossible for compact heat exchangers.

FIG. 3A depicts a schematic drawing of a second heat exchanger in accordance with the prior art. Heat exchanger 300 comprises tube 302, ruffled fins 304, inlet 306, outlet 308, and frame 310. Heat exchanger 300 is in accordance with a heat exchanger described by W. E. McCullough in U.S. Pat. No. 2,347,957, which was issued May 2, 1944, and which is incorporated by reference herein.

Tube 302 is a tubular-shaped conduit for conveying a first fluid along the z-direction as shown. Tube 302 is formed by flattening larger extruded tubing stock to form a central tube (i.e., tube 302) having attached flat projections.

These flat projections are run through rollers, or other suitable apparatus, to ruffle the projections into ruffled fins 304. As a result, ruffled fins 304 are wavy projections that extend laterally from tube 302.

FIG. 3B depicts a perspective view of a section of tube 302. Tube 302 comprises conduit 312, which conveys the first fluid along the z-direction. To form heat exchanger core 312, tube 302 is bent multiple times to form a continuous serpentine-shaped coil. In order to facilitate bending tube 302, portions of ruffled fins 304 are cut away near the regions of the bends. Heat exchanger core 312 is mounted into frame 310 such that each of the plurality of straight runs of tube 302 is oriented along the z-direction, and ruffled fins 304 project along the x-direction, as shown.

FIG. 3C depicts a schematic of a portion of heat exchanger 300. As evidenced by FIG. 3C, the bends of tube 302 are judiciously placed so that the crests 312 of ruffled fins 304 collectively define funnel-shaped “air scoops 314” that collectively form a plurality of “Venturi throats.” In other words, tube 302 is bent such that the wave of ruffled fins 304 of a first straight section is 180 degrees out of phase with the waves of adjacent straight sections. For example, referring to FIGS. 7 and 11 of the patent, McCullough discloses that for adjacent straight runs, the crests of the ruffled fins 304 are located “extending away from each other and beyond the width of the tube section 11. Due to this construction each of these ruffles acts as an air scoop and between them is an opening having a general funnel shape which narrows down to a flat opening 24 between the runs of tube section . . . ” And, “The cross section through these ruffles (FIG. 16) shows that the air stream first impinges forcefully against the fins 22 and 23, then rushes with increased velocity through the Venturi throat, 24, and expands to make contact with the ruffles 25 and 28.” See e.g., McCullough, page 2, second column, lines 3-20.

It is clear from this description that heat exchanger 300 is suitable only for exchanging heat between the first fluid and air. Further, heat exchanger 300 conveys the first fluid along the z-direction and the second fluid (i.e., air) along an orthogonal direction (i.e., the x-direction, as shown). As a result, the interaction length of heat two fluids in heat exchanger 300 is extremely short, which limits the efficiency of heat transfer. Still further, the 180 degree bends in the flow-path of the first fluid lead to a large pressure drop that reduces the efficiency of the heat exchanger and limits the size of such heat exchangers.

FIG. 4 depicts a perspective view of a heat exchanger in accordance with the illustrative embodiment of the present invention. Heat exchanger 110 comprises plates 402-1 and 402-2 and end plates 408-1 and 408-2. Although the illustrative embodiment comprises two plates 402, it will be clear to one skilled in the art, after reading this specification, how to specify, make, and use alternative embodiments of the present invention that comprise any practical number of plates 402. FIG. 4 is described with continuing reference to FIG. 1.

Heat exchanger 110 thermally couples working fluid 108 and seawater taken from a region of the water body 136. For example, heat exchanger 110-1 acts as an evaporator that heats working fluid 108 by transferring heat from warm seawater of surface region 118. In similar fashion, heat exchanger 110-2 acts as a condenser that cools vaporized working fluid 108 by thermally coupling it with cold seawater of deep water region 126.

Each of plates 402-1 and 402-2 (collectively referred to as plates 402) conveys seawater along the z-direction via conduits 404-1 and 404-2, respectively. Each of conduits 404-1 and 404-2 comprises a plurality of channels 406. In the illustrative embodiment, plates 402 are extruded plates of aluminum alloy. In some embodiments, plates 402 are made of another suitable material, such as aluminum, composite materials, graphite, graphite foam, and the like. It is preferable that plates 402 be made of a material that is substantially corrosion-free when exposed to seawater and/or common working fluids. Channels 406 are fluidically coupled to closed-loop conduit 106 via manifolds (not shown for clarity).

End plates 408-1 and 408-2 (collectively referred to as end plates 408) mate with plates 402-1 and 402-2, respectively. Plates 402 and end plates 408 collectively define a heat exchanger core that conveys both working fluid 108 and seawater along the z-direction while also fluidically isolating the fluids from one another.

Plates 402 are stacked together with end plates 408 to collectively define conduits 410-1, 410-2, and 410-3 (collectively referred to as conduits 410). Each of conduits 410 conveys working fluid 108 through heat exchanger 110 along the z-direction. Conduits 410 are described in more detail below and with respect to FIG. 6.

It should be noted that although the illustrative embodiment conveys seawater through conduits 404 and working fluid through conduits 410, in some embodiments seawater is conveyed through conduits 410 and working fluid is conveyed through conduits 404. In some alternative embodiments, a secondary fluid other than seawater is conveyed through the heat exchanger.

Plates 402 and end plates 408 are joined with a substantially galvanic corrosion-free joint, such as a friction-stir welding joint. In some embodiments, plates 402 and end plates 408 are joined using a different joining technology.

FIG. 5A depicts a top view of a portion of plate 402 prior to the formation of a wave-shape on its fins in accordance with the illustrative embodiment of the present invention. Plate 402 comprises body 502, straight fins 506, and partition 512. Plate 402 is representative of each of plates 402-1 and 402-2.

Body 502 is the central portion of plate 402, which comprises conduit 404. On each end (left and right ends, as depicted in FIG. 4), body 502 comprises a sidewall 508. Sidewall 508 projects above and below body 502 a distance necessary to provide a desired amount of clearance along the y-direction above fins 510 when plate 402 is joined with another plate 402 or an end plate 408. Sidewall 508 also provides a surface for joining plate 402 with another plate 402 or an end plate 408.

Straight fins 506 are projections that project from surface 504 of plate 402. Straight fins 506 are normal to surface 504.

Partition 512 is a straight wall that forms a sidewall for an end flow passage of conduit 410, as described below and with respect to FIGS. 6A and 6B.

Straight fins 506, sidewall 508, and body 502 are contiguous portions of a single extrusion that forms plate 402.

FIG. 5B depicts a top view of a portion of plate 402, after the formation of a wave-shape on its fins.

Straight fins 506 are formed into fins 510 through a conventional forming process, such as stamping, rolling, crimping, or hot forming. For example, a pair of stamping dies, having a desired wave shape, can be used to press this desired wave shape into straight fins 506. Alternatively, rollers or cams having a suitable forming surface can be rolled along the sides of straight fins 506 to deform them into fins 510. Such a rolling process offers an ability to form fins 510 in a continuous manner. Further parallel sets of rollers or cams enable the formation of a plurality of fins 510 at the same time.

Once straight fins 506 have been shaped, fins 510 are characterized by the desired wave shape along the z-direction, wherein the wave shape is that of a sinusoid having amplitude a and wavelength λ. In some embodiments, fins 510 are characterized by a periodic shape other than a sinusoid, such as triangular or chevron-shaped patterns.

FIG. 5C depicts a bottom view of a portion of plate 402, after the formation of a wave-shape on its fins. Plate 402 further comprises fins 516, which project from bottom surface 514 of body 502. Fins 516 are analogous to fins 510; however, fins 516 are formed such that, when plates 402-1 is flipped over and plates 402-1 and 402-2 are joined, the wave shapes of fins 510 and 514 nest such that there is a relative phase, φ, between them along the z-direction. In the illustrative embodiment, φ has a magnitude of approximately 180 degrees. In some embodiments, φ has a magnitude of approximately 90 degrees. In some embodiments, φ can have any magnitude within the range of 0 degrees to 360 degrees. The relationship of φ to the performance of heat exchanger 110 is discussed in more detail below and with respect to FIGS. 7 and 8A-C.

In similar fashion to conduit 410-2, conduit 410-1 is collectively defined by plate 402-1 and end plate 408-1. The top surface of end plate 408-1 is analogous to surface 504. Further, end plate 408-1 comprises fins 510 that project into conduit 410-1 from its top surface. The fins 510 of end plate 408-1 nest with fins 516 that project from surface 514 of plate 402-1 to form a plurality of flow passages in conduit 410-1. Still further, the top surface of end plate 408-2 is analogous to surface 514 and end plate 408-2 comprises fins 516 that project into conduit 410-3 from its bottom surface. The fins 516 of end plate 408-2 nest with fins 510 that project from surface 504 of plate 402-2 to form a plurality of flow passages in conduit 410-3. As a result, conduits 410-1 and 410-3 are analogous to conduit 410-2.

FIGS. 6A and 6B depict a representational cross-sectional and top view through line a-a, respectively, of conduit 410-2 in accordance with the illustrative embodiment of the present invention. When plates 402-1 and 402-2 are joined, fins 510 project from surface 504-1 into conduit 410-2. Fins 510 nest with fins 516 that project from surface 514-2 into conduit 410-2. As a result, fins 510 and 516 collectively define flow passages 602 within conduit 410-2. In addition, flow passage 604 is defined by the right-most (as depicted in FIG. 6A) fin 516 and partition 512, which projects from surface 504-1. In similar fashion, although not shown for clarity, left-most fin 510 and partition 512 that projects from surface 514-2 collectively define a second flow passage 604.

FIG. 6C depicts a detailed cross-sectional view of a portion of conduit 410-2. Each of fins 510 and 516 has substantially the same dimensions. While the specific dimensions of fins 510 and 516 are dependent upon the desired characteristics of heat exchanger 110, the dimensional relationships of fins 510 and 516 influence the manner in which working fluid 108 flows through conduit 410-2.

Each of fins 510 and 516 is characterized by a wave starting height of h1. The portion of each fin that comprises its wave shape is denoted as h2. In the illustrative embodiment, h2 is greater than h1 although it will be clear to one skilled in the art, after reading this specification, how to specify, make, and use alternative embodiments of the present invention wherein h2 is not larger than h1.

Fin-to-base clearance h3 denotes the clearance between the fins and surfaces 504 and 514 of conduit 410-2. The fin-to-base clearance is determined by the difference between the separation, s, between surfaces 504 and 514 and the sum of h1 and h2.

Each of the pluralities of fins 510 and 514 are characterized by the same fin-periodicity, d1. The wave-to-wave clearance, d2, is based on the fin-periodicity, d1, and the wave amplitude, a.

Judicious selection of h1, h2, h3, s, d1, d2, a, λ, and φ enables the design of conduits 410 that have high heat transfer efficiency.

It should be noted that prior art attempts to incorporate wavy channels (e.g., wavy fins) into conventional heat exchangers have utilized channels whose cross-sectional area remains constant along the direction of fluid flow. These channels are wavy in that they periodically abruptly change the flow direction of the fluid. Examples of such prior-art heat exchangers are found in “Forced Heat Convection in Wavy Fin Channel,” by Yang, et al., published in The Journal of Thermal Science and Technology, Vol. 3, pp. 342-354, (2008). A boundary layer, which forms in the flowing fluid, will separate and reattach at alternative periods of the wave thereby resulting in vortices around the flow channels. While these vortices improve the heat transfer coefficient of the heat exchanger, such heat exchangers are subject to a significant pressure drop (i.e., fluidic back pressure) through the length of the flow channels.

FIG. 7 provides operations of a method for thermally coupling a first fluid and a second fluid in accordance with the illustrative embodiment of the present invention. Method 700 begins with operation 701, wherein heat exchanger 110 is provided. FIG. 7 is described herein with continuing reference to FIGS. 1, and 4-6C and further reference to FIGS. 8A-C.

FIGS. 8A-C depict a flow passage 602-i within in conduit 410-2 at different points along the z-direction. The shape of flow passage 602-i is based on the relative orientations of shoulders 802 and wave portion 804 of fins 516-i and 510-i, respectively. Flow passage 602-i comprises sections 806, 808, and 810. The shape and cross-sectional area of section 808 changes along the length of conduit 410-2 and is defined by the configuration of fins 510-i and 516-i at each point along the z-axis.

At operation 702, conduits 404 are fluidically coupled to a region of water body 136. As a result, seawater is conveyed through heat exchanger 110 via conduits 404.

At operation 703, conduits 410 are fluidically coupled to working fluid 108. As a result, working fluid 108 is conveyed through heat exchanger 110 via conduits 410. Both the seawater and working fluid are conveyed along the z-direction, as described above; therefore, the interaction length between them is substantially the length of heat exchanger 110. A long interaction length enables a highly efficient transfer of thermal energy between the seawater and working fluid.

At operation 704, changes in the cross-sectional areas of sections 808 induce flow 812 of working fluid. These changes induce the flow of working fluid 108 between section 808 and sections 806 and 810 of each flow passage in the conduit.

FIG. 8A depicts flow passage 602-i at point A, when wave portions 804 of fins 510-i and 516-i are separated by a maximum distance. At this point along conduit 410-2, section 808 of flow passage 602-i has the largest cross-sectional area.

FIG. 8B depicts flow passage 602-i at point B, when fins 510-i and 516-i are straight. At this point along conduit 410-2, section 808 of flow passage 602-i has a smaller cross-sectional area than that depicted in FIG. 8A.

FIG. 8C depicts flow passage 602-i at point C, when fins 510-i and 516-i are separated by a minimum distance. At this point along conduit 410-2, section 808 of flow passage 602-i has the smallest cross-sectional area.

As working fluid 108 flows through conduit 410-2 from point A to point B to point C, the working fluid is squeezed from section 808 into each of sections 806 and 810. This creates turbulence in and around sections 806 and 810 due to flow 812. In some embodiments, flow 812 is a swirl flow. In some embodiments, flow 812 is a vortex flow. The turbulence in these sections significantly increases the thermal transfer efficiency of heat exchanger 110.

It should be noted that the combined cross-sectional area of passages 602 and 604 remains the same throughout the length of heat exchanger 110. This is readily seen by the fact that as the cross-section of section 808-i becomes smaller, its reduction in size is offset by a commensurate increase in the cross-sectional area of 808-i+1. The shift in cross-sectional area between neighboring flow passages is responsible for advantageously inducing turbulent flow in sections 806 and 810 of each flow passage. The conservation of the overall cross-sectional area of conduits 410 affords embodiments of the present invention additional advantage since the improved thermal transfer efficiency accrues without incurring an increase in flow resistance through the conduits. Still further, the lack of discontinuities in flow passages 602 and 604 enables heat exchanger 110 to avoid fluid flow interruptions inherent in prior-art heat exchangers, such as heat exchanger 200. Such fluid flow interruptions greatly increase fluid back pressure in these prior-art systems.

FIGS. 9A and 9B depict representational cross-sectional views through a portion of a heat exchanger, at two points along the direction of fluid flow, in accordance with an alternative embodiment of the present invention. Heat exchanger 900 is substantially identical to heat exchanger 400; however, in heat exchanger 900, plate 402-1 is replaced by plate 902-1.

Plate 902 is analogous to plate 402 prior to the formation of a wave-shape on its fins. In other words, plate 902 comprises fins 506, rather than fins 510, and a top view of plate 902 would be analogous to the top view of plate 402 depicted in FIG. 5A.

When plates 902-1 and 402-2 are joined as shown in FIGS. 9A and 9B, fins 506 project from surface 504-1 into conduit 410-2. Fins 506 nest with fins 516 that project from surface 514-2 into conduit 410-2. As a result, fins 506 and 516 collectively define flow passages 904 within conduit 410-2.

Each of flow passages 904 comprises sections 906 and 910, which are interposed by section 908. Sections 904, 906, and 910 are analogous to sections 806, 810, and 812, described above and with respect to FIGS. 8A-C.

In FIG. 9A, the cross-sectional view is taken at a point analogous to point A of the heat exchanger, as described above and with respect to FIG. 6B. In other words, the wave-shapes of fins 516 are depicted at their point of right-most deflection.

In FIG. 9B, the cross-sectional view is taken at a point analogous to point C of the heat exchanger, as described above and with respect to FIG. 6B. As a result, fins 516 as depicted in FIG. 9B are 180 degrees out of phase with those depicted in FIG. 9A.

As working fluid 108 flows through conduit 410-2 from point A to point C, the working fluid is squeezed from section 908 into each of sections 906 and 910. This creates turbulence in and around sections 906 and 910 due to flow into and out of section 908. In some embodiments, this flow is a swirl flow. In some embodiments, this flow is a vortex flow. As for heat exchanger 110, the turbulence in sections 906 and 910 significantly increases the thermal transfer efficiency of heat exchanger 900.

It should be noted that the combined cross-sectional area of passages 904, like that of passages 602, remains the same throughout the length of heat exchanger 110. This is readily seen by the fact that as the cross-section of section 908-i becomes smaller, its reduction in size is offset by a commensurate increase in the cross-sectional area of section 908-i+1.

It is to be understood that the disclosure teaches just one example of the illustrative embodiment and that many variations of the invention can easily be devised by those skilled in the art after reading this disclosure and that the scope of the present invention is to be determined by the following claims. 

1. A heat exchanger that thermally couples a first fluid and a second fluid, wherein the heat exchanger comprises: (1) a first plate comprising a first material that is thermally conductive, wherein the first plate comprises; (i) a first conduit for conveying the first fluid along a first direction, wherein the first conduit comprises at least one first channel; and (ii) a first plurality of first fins, wherein each first fin is continuous along the first direction, and wherein each first fin comprises a first fin portion that has a first periodic shape having a first periodicity along the first direction; and (2) a second plate comprising the first material, wherein the second plate comprises; (i) a second conduit for conveying the first fluid along the first direction, wherein the second conduit comprises at least one second channel; and (ii) a second plurality of second fins, wherein each second fin is continuous along the first direction; wherein the first plurality of first fins and the second plurality of second fins collectively define a plurality of passages for conveying the second fluid along the first direction, and wherein each of the plurality of passages comprises a first section, second section, and third section; and wherein the cross-sectional area of the first section varies along the first direction based on the first periodicity, and wherein the variation of the cross-sectional area of the first section induces flow of the second fluid between the first section and each of the second section and third section.
 2. The heat exchanger of claim 1 wherein each second fin is substantially straight along a second direction that is orthogonal with the first direction.
 3. The heat exchanger of claim 1 wherein each second fin comprises a second fin portion that has the first periodic shape having the first periodicity along the first direction.
 4. The heat exchanger of claim 3 wherein the first plurality of first fins and the second plurality of second fins are arranged along the first direction with a phase offset.
 5. The heat exchanger of claim 4 wherein the phase offset is 90 degrees.
 6. The heat exchanger of claim 4 wherein the phase offset is 180 degrees.
 7. The heat exchanger of claim 4 wherein the phase offset is 180 degrees.
 8. The heat exchanger of claim 1 wherein each first fin further comprises a third fin portion that is substantially straight along the first direction, and wherein the third fin portion interposes the first plate and the first fin portion.
 9. The heat exchanger of claim 8 wherein each second fin further comprises a fourth fin portion that is substantially straight along the first direction, and further wherein the fourth fin portion interposes the second plate and the second fin portion.
 10. The heat exchanger of claim 8 wherein the first fin portion has a first height and the third fin portion has a second height, and wherein the first height and the second height are substantially equal.
 11. The heat exchanger of claim 8 wherein the first fin portion has a first height and the third fin portion has a second height, and wherein the first height and the second height are unequal.
 12. The heat exchanger of claim 1 wherein the variation of the cross-sectional area of the first section induces turbulence in the second fluid in at least one of the second section and third section.
 13. The heat exchanger of claim 1 wherein the variation of the cross-sectional area of the first section induces a swirl flow of the second fluid in at least one of the second section and third section.
 14. The heat exchanger of claim 1 wherein the variation of the cross-sectional area of the first section induces a vortex flow of the second fluid in at least one of the second section and third section.
 15. A heat exchanger that thermally couples a first fluid and a second fluid, wherein the heat exchanger comprises: a first conduit for conveying the first fluid along a first direction; and a second conduit for conveying the second fluid along the first direction, wherein the second conduit comprises plurality of fins that are continuous along the first direction, and wherein each of the plurality of fins comprises a first fin portion that has a first periodic shape having a first periodicity along the first direction; wherein the plurality of fins collectively define a plurality of passages in the second conduit, and wherein each of the plurality of passages comprises a first section, second section, and third section, and wherein the first section has a cross-sectional area that varies along the first direction based on the first periodicity, and further wherein the variation of the cross-sectional area of the first section induces flow of the second fluid between the first section and each of the second section and third section.
 16. The heat exchanger of claim 12 wherein the variation of the cross-sectional area of the first section induces turbulence in at least one of the second section and third section.
 17. The heat exchanger of claim 12 wherein the variation of the cross-sectional area of the first section induces swirl flow in at least one of the second section and third section.
 18. The heat exchanger of claim 12 wherein the variation of the cross-sectional area of the first section induces vortex flow in at least one of the second section and third section.
 19. A method comprising: conveying a first fluid in a first conduit, wherein the first conduit conveys the first fluid along a first direction; conveying a second fluid in a plurality of passages, wherein the second fluid and first fluid are thermally coupled via the first conduit and the plurality of passages, and wherein each of the plurality of passages is continuous along the first direction and comprises a first section, second section, and third section, and wherein the cross-sectional area of the first section varies along the first direction based on the first periodicity; and inducing a first flow of the second fluid between the first section and each of the second section and third section, wherein the first flow is induced by the variation of the cross-sectional area of the first section along the first direction.
 20. The method of claim 19 wherein the first flow is induced such that a turbulent flow is induced in at least one of the second section and third section.
 21. The method of claim 19 wherein the first flow is induced such that a swirl flow is induced in at least one of the second section and third section.
 22. The method of claim 19 wherein the first flow is induced such that a vortex flow is induced in at least one of the second section and third section.
 23. The method of claim 19 further comprising providing the first conduit and second conduit, wherein providing the first conduit and second conduit are provided by operations comprising: providing a first plate, wherein the first plate comprises the first conduit and a first plurality of first fins; providing a second plate, wherein the first plate comprises a second conduit for conveying the first fluid along the first direction, and wherein the second conduit comprises at least one second channel, and further wherein the second plate comprises a second plurality of second fins; wherein the first plate and second plate are provided such that the first plurality of first fins and second plurality of second fins collectively define the plurality of passages. 